Hydraulic damper integrated into steering rack for attenuating steering nibble

ABSTRACT

The present invention provides a damping system for use with a mechanism such as a steering system wherein one element moves relative to another element. The system includes a variable volume chamber, a collection chamber, and an inertia track providing fluid communication between the two chambers. A fluid is disposed in the chambers and the inertia track. A compliant tuning member is in fluid communication with the variable volume chamber. The damping system is capable of attenuating vibration at chosen frequencies without adding undesired damping at other frequencies. In the preferred embodiment, the damping system is used to attenuate steering wheel torsional vibration without increasing low frequency damping that would deteriorate steering feel.

REFERENCE TO RELATED APPLICATIONS

This application is a continuation-in-part of U.S. patent applicationSer. No. 10/649,605, filed Aug. 28, 2003, which claims priority to U.S.Provisional Patent Application Ser. No. 60/480,219, filed Jun. 30, 2003.This application also claims priority to U.S. Provisional PatentApplication Ser. No. 60/630,160, filed Nov. 23, 2004, the entire contentof all of which are incorporated herein.

FIELD OF THE INVENTION

The present invention relates to steering arrangements and in particularto steering arrangements, which can provide tuned damping capable ofhandling large displacements.

BACKGROUND OF THE INVENTION

Existing vehicle wheel steering arrangements include a steering gear,tie rods, control arms, control arm bushings, and wheels. The frictionaldamping within the steering gear affects both steering feel, which isgenerally at frequencies less than 2 Hz, and the higher frequencychassis modes, which are typically in the range of 10-15 Hz.

Chassis modes can cause vibration to propagate through the steering gearand result in oscillations of the steering wheel, also referred to as“steering nibble”. Steering nibble is a significant issue for theautomotive industry.

Previous steering arrangements relied heavily on frictional dampingwithin the steering gear to attenuate steering nibble energy. Thisarrangement provides some chassis mode energy dissipation; however, theattenuation may be insufficient because frictional damping becomes lesseffective at higher frequencies.

Also, the automotive industry trend has been to reduce frictionaldamping, because decreasing 0-2 Hz damping positively affects customerperceived steering feel. Therefore, any arrangement that relies onincreased frictional damping to control 10-15 Hz vibrations willnegatively impact steering feel.

Some vehicle steering arrangements have employed a common viscousdamper, or dashpot, between the steering gear and tie rod, to adddamping to the system to attenuate steering nibble. Viscous devicesprovide damping across a broad frequency range, even though increaseddamping is only desired at the steering nibble frequency. Increaseddamping at low frequencies degrades steering feel, while increaseddamping at high frequencies transmits more vibration into the passengercompartment. The viscous damper provides this compromised performance aswell as a significant cost increase.

SUMMARY OF THE INVENTION

The present invention provides a variety of designs for damping systemsfor use with a variety of mechanisms or systems that require damping. Afirst preferred embodiment of the present invention provides a dampingsystem for use with a steering system of the type having a steering gearand a tie rod which moves toward and away from the steering gear tocause movement of a vehicle wheel. The damping system includes a bootdefining a working chamber. The boot has a first portion connected tothe steering gear and a second portion connected to the tie rod suchthat movement of the tie rod toward the steering gear reduces the volumeof the working chamber and movement of the tie rod away from thesteering gear increases the volume of the working chamber. An inertiatrack has a first end in fluid communication with the working chamber ofthe boot and an opposed second end. A collection chamber is in fluidcommunication with the second end of the inertia track. A fluid isdisposed in the working chamber, the inertia track, and at least aportion of the collection chamber. A compliant tuning member is in fluidcommunication with the working chamber. The tuning member is movable toaccommodate high frequency changes in the volume of the working chamber.

In some versions, the boot is a rolling boot and in some versions thecompliant tuning member is a diaphragm. The system may further include adiaphragm housing in fluid communication with the working chamber, withthe diaphragm forming a portion of the diaphragm housing. In additionalversions, the boot has a bulging stiffness and the diaphragm has adeflection stiffness that is substantially lower than the bulgingstiffness of the boot. The compliant tuning member may form at least apart of the boot. The compliant tuning portion of the boot may have astiffness substantially lower than the remainder of the boot, which maybe fiber reinforced. The collection chamber may be a chamber that is atleast partially filled with gas.

In further versions of the first embodiment, the steering gear is of thetype further having a second tie rod which moves towards and away fromthe steering gear. The system further includes a second boot defining asecond working chamber. The second boot has a first portion connected tothe steering gear and a second portion connected to the second tie rodsuch that movement of the tie rod toward the steering gear reduces thevolume of the second working chamber and movement of the tie rod awayfrom the steering gear increases the volume of the working chamber. Thesecond end of the inertia track is in fluid communication with thesecond boot such that the second boot forms the collection chamber.Alternatively, in a system including a second boot, a second inertiatrack and a second collection chamber may be provided, along with asecond compliant tuning member. The inertia tracks may have the samecross-sectional area in length, or the area and/or lengths may bedifferent from each other.

According to a further embodiment of the present invention, a dampingsystem is provided for use with a steering system of the type having asteering gear and a tie rod which moves toward and away from thesteering gear to cause movement of a vehicle wheel. The damping systemis tuned to at least partially damp a chosen frequency. The systemincludes a variable volume chamber in mechanical communication with thesteering gear and the tie rods such that movement of the tie rod in afirst direction with respect to the steering gear reduces the volume ofthe chamber and movement of the tie rod in a second direction withrespect to the steering gear increases the volume of the chamber. Thesystem further includes a collection chamber and an inertia track havingone end in fluid communication with the variable volume chamber and anopposed second end in fluid communication with the collection chamber. Afluid is disposed in the variable volume chamber, the inertia track andat least a portion of the collection chamber. A compliant tuning memberis in fluid communication with the chamber. The cross-sectional area andlength of the inertia track and the deflection stiffness of thecompliant tuning member are chosen such that the system damps movementof the tie rod relative to the steering gear generally at the chosenfrequency.

According to yet a further embodiment of the present invention, adamping system is provided for use with a mechanism having a firstelement which moves toward and away from a second element. The dampingsystem is tuned to at least partially damp a chosen frequency. Thesystem includes a variable volume chamber in mechanical communicationwith the mechanism such that movement of the first element in a firstdirection with respect to the second element reduces the volume of thechamber and movement of the first element in a second direction withrespect to the second element increases the volume of the chamber. Thesystem further includes a collection chamber and an inertia track havinga first end in fluid communication with the variable volume chamber andan opposed second end in fluid communication with the collectionchamber. A fluid is disposed in the variable volume chamber, the inertiatrack and at least a portion of the collection chamber. A volume offluid is displaced into or from the variable volume chamber by movementof the first element relative to the second element. The volumedisplaced per unit of movement is defined as dV/dx. An actual fluid massfor the inertia track is given by the formula: actual fluidmass=(cross-sectional area)×(length)×fluid density. The system has aneffective mass given by the formula: system effectivemass=((dV/dx)/(cross-sectional area))²×actual fluid mass. A complianttuning member is in fluid communication with the chamber. The complianttuning member has a deflection stiffness and an area. The system has asystem effective stiffness given by the formula: system effectivestiffness=((dV/dx)/(area of compliant member))²×deflection stiffness.The system has a resonant frequency given by the formula: systemresonant frequency=(system effective stiffness/system effectivemass)^(1/2). The cross-sectional area in length of the inertia track andthe deflection stiffness of the compliant tuning member are chosen suchthat the system resonant frequency is generally at the chosen frequency.Some versions of this embodiment are designed for use with a steeringsystem, wherein the first element is a tie rod and the second element isa steering rack. The system may further include a boot that defines thevariable volume chamber, with the boot having a first portion connectedto the steering rack and a second portion connected to the tie rod. Theboot may be a rolling boot. In some versions, the system effective massis at least 100 times the actual fluid mass, and in further versions thesystem effective mass is at least 200 times the actual fluid mass. Thesystem resonant frequency may be in the range of 10-25 hertz.

BRIEF DESCRIPTION OF THE DRAWINGS

Preferred embodiments of the invention are shown in the drawings,wherein:

FIG. 1 is a schematic of a vehicle steering arrangement having anintegrated hydraulically tuned damping system according to an embodimentof the invention;

FIG. 2 is a sectional view through one of the boots associated with theintegrated hydraulically tuned damping system of FIG. 1;

FIG. 3 is a sectional view through one of the boots associated with theintegrated hydraulically tuned damping system demonstrating thedistortion that yields its bulging stiffness;

FIG. 4 a is a sectional view through a boot of an alternate constructionhaving a separate diaphragm and associated with the integratedhydraulically tuned damping system;

FIG. 4 b is a sectional view through a fiber-reinforced boot andintegrated diaphragm;

FIG. 5 a is a schematic of a vehicle steering arrangement having anintegrated hydraulically tuned damping system of an alternateconstruction, where each boot is connected to its own separate inertiatrack and fluid collection chamber;

FIG. 5 b is a sectional view through a boot of an alternate constructionwhere the boot is connected to its own separate inertia track and fluidcollection chamber;

FIG. 5 c is a schematic of a vehicle steering arrangement having anintegrated hydraulically tuned damping system of an alternateconstruction, where both boots are connected to a common auxiliary fluidcollection unit;

FIG. 5 d is a sectional view through a version of a common auxiliaryfluid collection unit for use with the present invention;

FIGS. 6 a shows the frequency damping characteristics of a steering gearwith low internal frictional damping;

FIGS. 6 b shows the frequency damping characteristics of a steering gearwith high internal frictional damping;

FIGS. 6 c shows the frequency damping characteristics of a hydraulicallytuned damping system integral to the steering gear;

FIG. 7 is a schematic of a vehicle steering arrangement in which ahydraulically tuned damping system has been integrated within thesteering gear housing;

FIG. 8 is a schematic of the stand-alone hydraulically tuned dampingsystem, which can be used for general application to attenuatefrequency-specific vibration;

FIG. 9 is a schematic of a stand-alone piston/cylinder typehydraulically tuned damping system, which can be used for generalapplication to attenuate frequency-specific vibration;

FIG. 10 is a schematic of a boot type hydraulically tuned damping systemhaving multiple boots, where a second opposed boot acts as the fluidcollection chamber for the first boot; and

FIG. 11 is a schematic of a piston/cylinder type hydraulically tuneddamping system having multiple boots, where a second cylinder acts asthe fluid collection chamber for the first cylinder.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A modified vehicle steering suspension arrangement is shownschematically in FIG. 1. This steering arrangement includes ahydraulically tuned steering damping system, which is comprised of tworolling boots 2 connected by an inertia track 4 and filled with workingfluid 6. The hydraulically tuned steering damping system, and the newrolling boots 2 in particular, are shown in greater detail in FIG. 2.

The two rolling boots 2 replace the common dust boots found on steeringgears, whose function was to prevent dirt from entering the steeringgear 7. The inertia track 4 may replace the common air equalization tubefound on steering gears, whose function is to prevent the air trappedwith the dust boots from building up pressure.

The hydraulically tuned steering damping system works on some similarprinciples as existing hydromounts, however, it is designed toaccommodate much greater displacements than hydromounts and does not addsignificant static stiffness to the systems as hydromounts do. Therolling boots 2 accommodate displacements greater than 25 mm withoutadding significant low frequency stiffness to the system.

The steering system is designed such that desired tie-rod 10 lateraldisplacements are caused by driver-controlled rotational steering wheelinputs. Similarly, undesirable energy can be transmitted in the reversedirection: from chassis resonances into the tie-rods 10, ultimatelyresulting in a rotation of the steering wheel.

When the tie-rod 10 is laterally displaced, it causes one end of theboot 2 to move with it because they are clamped 3 together. As the endof the boot 2 moves inward toward the steering gear 7, it rolls furtherunder itself, making the working chamber 5 smaller in volume. Thisvolumetric change applies pressure to the working fluid 6 and causesmovement of the working fluid 6 through the inertia track 4.

Thus, the working fluid 6 is forced to move through the inertia track 4between the contracting working volume 5 and into the expanding volumewithin the boot on the opposite side of the steering gear. The totalcombined volume of the two boots is constant; therefore there is nosignificant pressure increase at low frequencies. For this case, theworking fluid 6 is free to move through the inertia track 4 with lowresistance. In this embodiment, each boot serves as a collection chamberfor the opposite boot.

The inertia track 4 connects the contracting working volume 5 with theexpanding volume on the opposite side. Structurally, the inertia trackis a tube or passage that allows the flow of fluid there through. Thelength of the inertia track 4 and the size thereof contribute to thetuning and effectiveness of the hydraulically tuned steering dampingsystem. The inertia track 4 is designed such that the mass of the fluidin the inertia track is scaled up to act as a large effective mass. Thismass scaling effect is caused by the impedance of the fluid that mustflow from the relatively large working chamber's cross-section into thesmall nozzle-like inertia track's cross-section.

FIG. 3 shows the distortion 12 that occurs in the boot 2 when theworking fluid 6 is not free to flow through the inertia track 4. Thisdistortion 12 is defined by a bulging stiffness and effectivelyfunctions as a deformable diaphragm, similarly to how the bulgingstiffness of a traditional hydromount functions as a deformablediaphragm. The portion of the boot that acts as a deformable diaphragmin this embodiment may also be referred to as a compliant tuning member.As will be clear to those of skill in the art, the compliant tuningmember is in fluid communication with the working chamber 5. Thediaphragm or compliant tuning member may take other forms, some of whichwill be described later.

It is the combination of the movement of fluid through the inertia track4 and the bulging distortion 12 (which defines an analogous diaphragm)that provide the tunable response characteristics of this embodiment ofa hydraulically tuned steering damping system. The resonant frequency ofany single degree of freedom tuned mass absorber is a simple function ofthe absorber mass and the stiffness of the spring on which it ismounted. The hydraulically tuned damping system's resonant frequency issimilarly a function of the scaled effective fluid mass and thecompliant member's stiffness. The scaled effective fluid mass alsoaffects the total output damping of the damping system at the resonantfrequency. Below the resonant frequency, output damping is low becausefluid is free to flow through the inertia track 4. Above the resonantfrequency, output damping is low because the compliant tuning member 12is compliant for the lower vibration amplitudes seen at highfrequencies.

The scaled effective fluid mass is one of the primary designconsiderations that affect the amount of damping achieved at theresonant frequency. The effective mass is increased when the workingchamber's 5 cross sectional area is increased or when the inertiatrack's 4 cross sectional area is decreased. Also, since a longerinertia track 4 causes more fluid mass to be in the inertia track 4,increasing inertia track 4 length is another design method to increasethe total effective mass and therefore the system's output damping.

As an example, a hydraulically tuned damping system achieved 10 Ns/mm ofoutput damping at 15 Hz. To achieve this level of damping the scaledeffective fluid mass is approximately 40 Kg. One way to achieve thiseffective mass would be to have a working chamber cross-section area of2000 mm², an inertia track that has cross section area of 70 mm² andlength of 80 mm. Of course, the compliant member's stiffness affects thefrequency at which this damping peak would occur, and to tune a 40 Kgmass to resonate near 15 Hz, the system's stiffness is approximately 400N/mm.

These damping characteristics differ drastically from the traditionalsteering arrangements' non-tunable damping, which is caused by frictionbetween the stationary steering gear housing 7 and the internalcomponents that move together with the tie-rods 10. Increased frictionaldamping can help alleviate 10-15 Hz lateral tie-rod vibration energycausing nibble, but increased friction is more influential in increasingdamping in the 0-2 Hz range, which the driver perceives as poor steeringfeel.

Returning to FIG. 2, it can be appreciated that the boot 2 behaves verydifferently depending on whether the working fluid 6 is allowed to flowfreely through the inertia track 4 or not. If the fluid 6 is free tomove, required volumetric changes are easily accommodated by the boot's2 rolling action without any significant stiffness effect. If the fluid6 is trapped in the inertia track 4, volumetric changes cannot be easilyaccommodated and the boot 2 must bulge 12, with significant stiffnesseffect upon the system.

FIG. 4 a shows an alternate hydraulically tuned steering damping system,which has a separate deformable diaphragm 16 attached to the structurethat houses the boot 2. In this system, the bulging stiffness of theboot 2 is designed orders of magnitude higher than the separatedeformable diaphragm 16, so that the compliance of the diaphragm 16controls the bulging stiffness used to tune the damping frequency of thesystem. For example, a typical diaphragm stiffness would be 100 N/mm. Byreinforcing the boot material 17 with a fiber mesh 18 as shown in FIG. 4b, the boot 2 can achieve a bulging stiffness of over 1000 N/mm. Thetotal bulging stiffness of these two components together would bedominated by the most compliant member, the diaphragm. Therefore, inthis example, the combined bulging stiffness would still be close to 100N/mm and the reinforced boot's bulging stiffness would have littlestiffness effect. The deformable diaphragm of this embodiment may besaid to form part of a diaphragm housing, which is in fluidcommunication with the working chamber 5 of the rolling boot 2.

FIG. 4 b also shows that a diaphragm can be designed into the bootitself by not reinforcing a desired material section 20 with fiber. Thisnon-reinforced section 20 would be more compliant than thefiber-reinforced boot 18 and could also be designed with a thinner wallsection than the rest of the boot. Similarly to the previously describedexternal diaphragm, this integrated diaphragm 20 could be designed tohave a stiffness of a typical diaphragm, approximately 100 N/mm, andwould be the controlling system compliance when paired with a reinforcedboot's bulging stiffness of over 1000 N/mm.

It should also be noted that the present invention may utilize boots orbellows with configurations different than the preferred rolling boot.For example, a more traditional pleated bellows or boot may be used insome applications.

FIGS. 5 a and 5 b show another alternate hydraulically tuned steeringdamping system, which has separate inertia tracks 4 for each boot andexternal fluid collection chambers 24 for each inertia track. Eachcollection chamber 24 is partially filled with a low pressure gas 26which accommodates fluid 6 volumetric changes and functions similarly tothe bellows of a traditional hydromount. Each inertia track 4 can betuned to attenuate an individual frequency or they can be tuned toattenuate the same frequency. This system can use either the boot'sbulging stiffness 12 or a separate deformable diaphragm, such as 16 or20, to achieve the required bulging stiffness. In the version of thesystem shown in FIG. 1, each boot 2 serves as a collection chamber forthe opposite boot. In FIGS. 5 a and 5 b, the external collectionchambers 24 serve this purpose.

FIGS. 5 c and 5 d show another alternate hydraulically tuned steeringdamping system, which has separate inertia tracks 4 for each boot, butone collection chamber 32. Each inertia track 4 can be tuned toattenuate an individual frequency or they can be tuned to attenuate thesame frequency. This system can be used with either the boot' bulgingstiffness 12 or a separate deformable diaphragm 16 or 20 to achieve therequired bulging stiffness.

FIGS. 6 a through 6 c show different systems and their damping vs.frequency characteristics. In each of the diagrams, different frequencyranges are shown which are particular concerns. The low frequency rangecontrols the steering feel characteristics, typically from 0-2 Hz, withthe chassis modes causing steering nibble shown in a band of about 10-15Hz. The next band of interest is much higher and has to do with highfrequency excitation and this is found at the right end of the graphs.

FIG. 6 a shows the performance of a steering gear designed with lowfrictional damping. With this system, there is little damping at lowfrequencies to disturb steering feel, but also little damping toattenuate steering nibble.

FIG. 6 b shows the performance of a steering gear designed with highfrictional damping. With this system, there is sufficient damping tocontrol steering nibble, but too much damping at low frequencies,thereby causing degraded steering feel.

FIG. 6 c shows the performance of a hydraulically tuned steering dampingsystem that is designed to have peak damping for the steering nibblemode. As can be seen, the structure provides minimal damping in the lowfrequency range and therefore has minimal impact on steering feel.

FIG. 7 shows an alternate hydraulically tuned steering damping system,which is integrated inside the structure of the steering gear, therebypotentially realizing additional cost savings. The system contains dustboots 42, an air equalization tube 43, a pinion 49, and a steering gearhousing 47 and its internal components. The steering gear contains apiston 40 that slides with the rack 50. Bearings 51 constrain the rack50 within the gear housing 47 and fluid is sealed. A piston 40 definestwo variable volume chambers 45 within the gear housing 47 and moves theworking fluid 46 through the inertia track 44 and between the volumechambers 45 when there is relative motion between the tie rod 41 and thesteering gear housing 47. Compliant tuning members, in the form ofdiaphragms 48, are integrated into the gear housing and are in fluidcommunication with the volume chambers 45. The compliance of thediaphragm 48 controls the bulging stiffness used to tune the dampingfrequency of the system. As will be clear to those of skill in the art,the integrated system of FIG. 7 may take other forms than illustrated.

FIG. 8 shows a stand-alone hydraulically tuned damping system, which canbe used to attenuate vibration for other general applications. A rollingboot 2 attached to a housing 9 defines a variable volume chamber 5,which is filled with a working fluid 6. The working fluid is displacedthrough an inertia track 4 and into a secondary collection chamber 36. Alow pressure gas 26 inside of the collection chamber 36 accommodatesfluid 6 volumetric changes. This system can use either the boot'sbulging stiffness 12 or a separate deformable diaphragm 16 or 20 toachieve the required bulging stiffness.

It is also possible with these designs to employ technologies found inexisting hydromounts such as multiple inertia tracks, floatingdiaphragms and other amplitude decoupling devices, high dampeddiaphragms, etc., which allow the designer to tailor the stiffnessfrequency curve of the damper.

Some embodiments of the hydraulically tuned steering damping systemutilize a rolling boot 2 design rather than the traditional moldedrubber shoulder design used in existing hydromount technology. Thisdesign differs functionally from the existing technology in that themain rubber element of a traditional hydromount provides significantstiffness to the system whether or not fluid is allowed to flow freelythrough the inertia track. At high frequencies when fluid is trapped inthe inertia track, the main rubber element provides a bulging stiffness.At low frequencies, when fluid flows freely through the inertia track,the traditional hydromount's main spring still provides a significantstatic stiffness. For example, a typical hydromount may provide 100 N/mmor more of static stiffness. In the present design, the rolling boot 2provides a significant stiffness only during high frequency bulging. Atlow frequencies, the rolling boot 2 rolls upon itself and addsnegligible stiffness to the system. Preferably, the static stiffness isof the rolling boot is less than 50 N/mm, and more preferably the staticstiffness is less than 10 N/mm and most preferably it is less than 1N/mm. For the purposes of this application, the low frequency stiffnessof the rolling boots, or the static stiffness, may be referred to as thelow frequency or static stiffness of the variable volume chamber. Thevariable volume's static stiffness and the static stiffness effect ofthe low-pressure gas, which is present in some embodiments, combine todefine the total static stiffness of the whole damping system. This isthe total resistance to movement experienced by the tie rod due to thedamping system at frequencies substantially below the tuned dampingfrequency. As will clear to those of skill in the art, this lowfrequency or static stiffness experienced by the tie rod due to thedamping system is very small because the damping system does not resistthe flow of fluid through the inertia track. Preferably, the totalstatic stiffness or low frequency stiffness of the entire damping systemis also less than 50 N/mm, and more preferably less than 10 N/mm, andmay be as low as 1 N/mm or less. Additionally, the damping system'sboots can undergo large strains with minimal rubber stress and thuscause negligible static stiffness increase. A hydromount's main spring,however, cannot undergo large strains without experiencing significantrubber stress and hence significant static stiffness effect.

Similarly, a damping system embodiment employing a piston/cylinder wouldhave no rubber stress when the piston is stroked and hence would have nosignificant static stiffness. The static stiffness of a piston/cylindersystem would be preferably under 20 N/mm, more preferably 10 N/mm, andmost preferably under 1 N/mm.

The rolling boot 2 design of the hydraulically tuned steering dampingsystem also differs functionally from an existing hydromount in that themain rubber element of a traditional hydromount cannot accommodate largedeflections above 20 mm, and works optimally at deflections under 10 mm.The rolling boot 2 easily accommodates deflections of over 25 mm, andpreferably over 50 mm, and more preferably above 100 mm. As will beclear to those of skill in the art, the maximum deflection is onlylimited by the boot travel, or the travel of an external cylinder, inembodiments using a cylinder to define a variable volume chamber orworking chamber. Typical maximum deflections are 100 mm, but could be ashigh as 250 mm or even as high as 2000 mm for some applications.

The working fluid 6 used in current hydromounts can be used in thehydraulically tuned steering damping system application. Other fluidsmay also be utilized which have different properties and lubricationsthrough viscosity, density, etc. Power steering fluid, which is alreadyused internally in the steering gear and is readily available insteering gear manufacturing facilities, presents another logical choiceof working fluid 6.

Various forms of inertia track 4 have been shown. All of the inertiatracks have by definition, an area of A_(i) and a length L_(i). Thecompliant tuning member caused by the bulging of the boot 12, by anexternal diaphragm 16, or by integrating a diaphragm into the boot 20,has by definition, an area Ad and is made of a compliant material. Thecompliant member does not need to be capable of handling large fluiddisplacements, since large displacements occur at relatively lowfrequencies where the fluid 6 is able to move through inertia track 4 ata rate that is sufficient to prevent excessive pressure build ups.

In the hydraulically tuned steering damping system it can be appreciatedthe amount of fluid 6 displaced from the collapsing working chamber 5 isequal to the area of an equivalent analogous piston times the distancethat the piston moves. The hydraulic fluid is assumed to beincompressible.

At low frequencies, the displaced fluid volume flows through the inertiatrack and into the expanding collection volume formed by the other boot,or into a separate collection chamber. This occurs for large amplitudedisplacements as well as small amplitude displacements.

The cross-sectional area of the analogous piston is large with respectto the cross sectional area of the inertia track. Therefore, a unitdisplacement of the boot/piston requires a much larger displacement offluid through the inertia track. The movement of fluid through theinertia track is increased with respect to the movement of theboot/piston. This scaling effect makes the few grams of the fluid in theinertia track appear to have a mass of many hundreds or thousands timeslarger. The gain is equal to:system effective mass=((Area of Piston)/(Area of Inertia Track))²×actualfluid masswhere actual fluid mass=fluid density×Area of Inertia Track×InertiaTrack Length

A more generic way to express this equation issystem effective mass=((dV/dx)/(Area of Inertia Track))²×actual fluidmass

where dV/dx is the volume of fluid displaced from the working chamberper unit of length the boot is collapsed. The scaling from the actualfluid mass to the system effective mass is typically over 200, and oftenover 2000.

At high frequencies, the inertial effects become quite large and theacceleration and therefore the displacement of the fluid in the inertiatrack approaches zero. A flexible diaphragm, such as caused by theboot's bulging stiffness, allows the effective mass in the inertia trackto decouple from the moving piston. The volume change that accompaniesthe piston movement is taken up by deflection of the diaphragm. Thedeflecting diaphragm adds high frequency stiffness to the system. Thearea of the diaphragm may or may not be equivalent to the area of thepiston or analogous piston, so that there may be a scaling effect asthere was with the inertia track. The diaphragm introduces a gain thatis equal tosystem effective stiffness=((Area of Piston)/(Area ofDiaphragm))²×diaphragm stiffness

Again, a more generic way to express this equation issystem effective stiffness=((dV/dx)/(Area of Diaphragm))²×diaphragmstiffnesswhere dV/dx is the volume of fluid displaced from the working chamberper unit of length the boot is collapsed.

The diaphragm can only accommodate small volume changes. This is not anissue since at high frequencies, where diaphragm motion is necessary,the displacements are quite small. At low frequencies wheredisplacements are high, the fluid moves through the inertia track andthe diaphragm is not significantly deflected.

At a certain frequency, the effective mass of the fluid in the inertiatrack resonates on the effective stiffness of the diaphragm, thusachieving significant output motion even when excited with little inputmotion. This resonant frequency is designed to occur at or near thevehicle's steering nibble frequency. As the effective mass resonates, ittransitions from in-phase motion to out-of-phase motion. During thistransition, the effective mass oscillates with an approximate 90-degreephase lag with respect to the excitation. This generates an enormousamount of effective damping. This damping can be used to attenuate thechassis mode of interest.

The resonant frequency for the system is given by the formula:system resonant frequency=(system effective stiffness/system effectivemass)^(1/2)

The cross sectional area and length of the inertia track and thedeflection stiffness of the diaphragm are chosen such that the systemresonant frequency is generally at the chosen damping frequency. Theresonant frequency for a steering system may be in the range of 10-25Hz, or more preferably in a range of 10-15 Hz, though a damping systemaccording to the present invention may be tuned to damp otherfrequencies.

It can be appreciated that the type of bellows used in a traditionalhydromount will not be adequate to accommodate the large volumetricchanges experienced in a long stroke hydraulic tuned damper. This istrue whether the long stroke hydraulically tuned damper is of thepiston/cylinder type shown in FIG. 9 or boot type shown in FIG. 8. Inthe design of FIG. 9, the piston/cylinder defines the variable volumechamber. This approach provides a long stroke and low static stiffnesssimilar to the boot version discussed earlier. The static stiffness ofthe piston/cylinder is less than 20 N/mm, and more preferably the staticstiffness is less than 10 N/mm and most preferably it is less than 1N/mm. For the purposes of this application, the stiffness of thepiston/cylinder at low frequencies, or the static stiffness, may bereferred to as the low frequency or static stiffness of the variablevolume chamber. The variable volume's static stiffness and the staticstiffness effect of the low-pressure gas, which is present in someembodiments, combine to define the total static stiffness of the wholedamping system. This is the resistance to movement experienced by thetie rod due to the damping system at frequencies substantially below thetuned damping frequency. The piston/cylinder easily accommodatesdeflections of over 25 mm, and preferably over 50 mm, and morepreferably above 100 mm. As will be clear to those of skill in the art,the maximum deflection is only limited by the piston travel. Typicalmaximum deflections are 100 mm, but could be as high as 250 mm or evenas high as 2000 mm for some applications.

One method of accommodating large volumetric changes without addingsignificant stiffness is to partially fill the fluid collection chamberwith a compressible low-pressure gas 27 and 26.

Another method is to attach two long stroke hydraulically tuned dampersin an opposing arrangement, as shown in FIGS. 10 and 11. The two devicesare arranged such that the volume of collapse in the first chamber isequal to the volume of expansion in the second chamber when the systemis given a linear displacement input. Therefore, each device functionsas a collection chamber for the other.

Long stroke hydraulically tuned dampers can also be used forapplications other than steering nibble. One example application is fordamping of brake roughness. By placing hydraulically tuned dampersaccording to the present invention at the lower control arm bushinglocations, the longitudinal vibration energy from the brake system canbe attenuated. Long stroke hydraulically tuned dampers for otherapplications are similar to the above described embodiments in that theyinclude a variable volume chamber interconnected with the mechanism, aninertia track connected to the chamber, a collection chamber connectedto the other end of the inertia track, and a compliant tuning member ordiaphragm in fluid communication with the variable volume chamber. Thedampers are designed in accordance with the above teachings, and may beused to damp a variety of frequencies.

Although various preferred embodiments of the present invention havebeen described herein in detail, it will be appreciated by those skilledin the art that variations may be made thereto without departing fromthe spirit of the invention or the scope of the appended claims.

1. A damping system for use with a steering system of the type having asteering gear and a tie rod which moves toward and away from thesteering gear to cause movement of a vehicle wheel, the systemcomprising: a boot defining a working chamber, the boot having a firstportion connected to the steering gear and a second portion connected tothe tie rod such that movement of the tie rod toward the steering gearreduces the volume of the working chamber and movement of the tie rodaway from the steering gear increases the volume of the working chamber;an inertia track having a first end in fluid communication with theworking chamber of the boot and an opposed second end; a collectionchamber in fluid communication with the second end of the inertia track;a fluid disposed in the working chamber, the inertia track, and at leasta portion of the collection chamber; and a compliant tuning member influid communication with the working chamber, the tuning member beingmovable to accommodate high frequency changes in the volume of theworking chamber.
 2. The damping system according to claim 1, wherein theboot is a rolling boot.
 3. The damping system according to claim 1,wherein the compliant tuning member is a diaphragm.
 4. The dampingsystem according to claim 3, further comprising a diaphragm housing influid communication with the working chamber, the diaphragm forming aportion of the diaphragm housing.
 5. The damping system according toclaim 4, wherein the boot has a bulging stiffness and the diaphragm hasa deflection stiffness substantially lower than the bulging stiffness ofthe boot.
 6. The damping system according to claim 1, wherein thecompliant tuning member forms at least a part of the boot.
 7. Thedamping system according to claim 6, wherein the compliant tuning partof the boot has a stiffness substantially lower than the remainder ofthe boot.
 8. The damping system according to claim 7, wherein at least aportion of the remainder of the boot is fiber reinforced.
 9. The dampingsystem according to claim 1, wherein the collection chamber comprises achamber at least partially filled with a gas.
 10. The damping systemaccording to claim 1, wherein the steering gear is of the type furtherhaving a second tie rod which moves towards and away from the steeringgear, the system further comprising: a second boot defining a secondworking chamber, the second boot having a first portion connected to thesteering gear and a second portion connected to the second tie rod suchthat movement of the tie rod toward the steering gear reduces the volumeof the second working chamber and movement of the tie rod away from thesteering gear increases the volume of the working chamber; the secondend of the inertia track being in fluid communication with the secondboot, such that the second boot forms the collection chamber.
 11. Thedamping system according to claim 1, wherein the steering gear is of thetype further having a second tie rod which moves towards and away fromthe steering gear, the system further comprising: a second boot defininga second working chamber, the second boot having a first portionconnected to the steering gear and a second portion connected to thesecond tie rod such that movement of the tie rod toward the steeringgear reduces the volume of the second working chamber and movement ofthe tie rod away from the steering gear increases the volume of theworking chamber; a second inertia track having a first end in fluidcommunication with the second working chamber and an opposed second end;a second collection chamber in fluid communication with the second endof the second inertia track; a second compliant tuning member in fluidcommunication with the second working chamber; and a fluid disposed inthe second working chamber, the second inertia track, and at least aportion of the second collection chamber.
 12. The damping systemaccording to claim 11, wherein the inertia tracks each have a crosssectional area and a length, the area and length of the tracks being thesame.
 13. The damping system according to claim 11, wherein the inertiatracks each have a cross sectional area and a length, the area and/orlengths of the tracks being different from each other.
 14. A dampingsystem for use with a steering system of the type having a steering gearand a tie rod which moves toward and away from the steering gear tocause movement of a vehicle wheel, the damping system being tuned to atleast partially damp a chosen frequency, the system comprising: avariable volume chamber in mechanical communication with the steeringgear and the tie rod such that movement of the tie rod in a firstdirection with respect to the steering gear reduces the volume of thechamber and movement of the tie rod in an opposite second direction withrespect to the steering gear increases the volume of the chamber; acollection chamber; an inertia track having a first end in fluidcommunication with the variable volume chamber and an opposed second endin fluid communication with the collection chamber, the inertia trackhaving a cross sectional area and a length; a fluid disposed in thevariable volume chamber, the inertia track and at least a portion of thecollection chamber; and a compliant tuning member in fluid communicationwith the chamber, the flexible tuning member having a deflectionstiffness; the cross sectional area and length of the inertia track andthe deflection stiffness of the compliant tuning member being chosensuch that the system damps movement of the tie rod relative to thesteering gear generally at the chosen frequency.
 15. A damping systemfor use with a mechanism having a first element which moves toward andaway from a second element, the damping system being tuned to at leastpartially damp a chosen frequency, the system comprising: a variablevolume chamber in mechanical communication with the mechanism such thatmovement of the first element in a first direction with respect to thesecond element reduces the volume of the chamber and movement of thefirst element in an opposite second direction with respect to the secondelement increases the volume of the chamber, the variable volume chamberhaving a stroke sufficient to allow displacements of the first elementrelative to the second element of at least 100 mm; a collection chamber;an inertia track having a first end in fluid communication with thevariable volume chamber and an opposed second end in fluid communicationwith the collection chamber, the inertia track having a cross sectionalarea and a length; a fluid disposed in the variable volume chamber, theinertia track and at least a portion of the collection chamber, thefluid having a fluid density; a volume of fluid being displaced into orfrom the variable volume chamber by movement of the first elementrelative to the second element, the volume displaced per unit ofmovement being defined as dV/dx; an actual fluid mass for the inertiatrack being given by the formula;actual fluid mass=(cross sectional area)×(length)×fluid density; thesystem having a system effective mass given by the formula;system effective mass=((dV/dx)/(cross sectional area))×actual fluidmass; a compliant tuning member in fluid communication with the chamber,the compliant tuning member having a deflection stiffness and an area;the system having a system effective stiffness given by the formula;system effective stiffness=((dV/dx)/(Area of compliantmember))²×deflection stiffness; the system having a resonant frequencygiven by the formula;system resonant frequency=(system effective stiffness/system effectivemass)^(1/2); the cross sectional area and length of the inertia trackand the deflection stiffness of the compliant tuning member being chosensuch that the system resonant frequency is generally at the chosenfrequency.
 16. The damping system according to claim 15, wherein themechanism is a steering system, the first element is a tie rod, and thesecond element is a steering gear housing.
 17. The damping systemaccording to claim 16, further comprising a boot defining the variablevolume chamber, the boot having a first portion connected to thesteering gear housing and a second portion connected to the tie rod. 18.The damping system according to claim 17, wherein the boot is a rollingboot.
 19. The damping system according to claim 15, wherein the systemeffective mass is at least 100 times the actual fluid mass.
 20. Thedamping system according to claim 19, wherein the system effective massis at least 200 times the actual fluid mass.
 21. The damping systemaccording to claim 15, wherein the system resonant frequency is in therange of 10-25 Hz.
 22. A damping system for use with a mechanism havinga first element which moves toward and away from a second element, thedamping system being tuned to at least partially damp a chosenfrequency, the system comprising: a variable volume chamber inmechanical communication with the mechanism such that movement of thefirst element in a first direction with respect to the second elementreduces the volume of the chamber and movement of the first element inan opposite second direction with respect to the second elementincreases the volume of the chamber; a collection chamber; an inertiatrack having a first end in fluid communication with the variable volumechamber and an opposed second end in fluid communication with thecollection chamber, the inertia track having a cross sectional area anda length; a fluid disposed in the variable volume chamber, the inertiatrack and at least a portion of the collection chamber, the fluid havinga fluid density; a gas disposed in at least a portion of the collectionchamber; a volume of fluid being displaced into or from the variablevolume chamber by movement of the first element relative to the secondelement whereby the gas disposed in the collection chamber is compressedand expanded, the volume of fluid displaced per unit of movement beingdefined as dV/dx; an actual fluid mass for the inertia track being givenby the formula;actual fluid mass=(cross sectional area)×(length)×fluid density; thesystem having a system effective mass given by the formula;system effective mass=((dV/dx)/(cross sectional area))²×actual fluidmass; a compliant tuning member in fluid communication with the chamber,the compliant tuning member having a deflection stiffness and an area;the system having a system effective stiffness given by the formula;system effective stiffness=((dV/dx)/(Area of compliantmember))²×deflection stiffness; the system having a resonant frequencygiven by the formula;system resonant frequency=(system effective stiffness/system effectivemass)^(1/2); the cross sectional area and length of the inertia trackand the deflection stiffness of the compliant tuning member being chosensuch that the system resonant frequency is generally at the chosenfrequency.
 23. A damping system for use with a mechanism having a firstelement which moves toward and away from a second element, the dampingsystem being tuned to at least partially damp a chosen frequency, thesystem comprising: a variable volume chamber in mechanical communicationwith the mechanism such that movement of the first element in a firstdirection with respect to the second element reduces the volume of thechamber and movement of the first element in an opposite seconddirection with respect to the second element increases the volume of thechamber; a collection chamber; an inertia track having a first end influid communication with the variable volume chamber and an opposedsecond end in fluid communication with the collection chamber, theinertia track having a cross sectional area and a length; a fluiddisposed in the variable volume chamber, the inertia track and at leasta portion of the collection chamber, the fluid having a fluid density; avolume of fluid being displaced into or from the variable volume chamberby movement of the first element relative to the second element, thevolume displaced per unit of movement being defined as dV/dx; an actualfluid mass for the inertia track being given by the formula;actual fluid mass=(cross sectional area)×(length)×fluid density; thesystem having a system effective mass given by the formula;system effective mass=((dV/dx)/(cross sectional area))²×actual fluidmass; a compliant tuning member in fluid communication with the chamber,the compliant tuning member having a deflection stiffness and an area;the system having a system effective stiffness given by the formula;system effective stiffness=((dV/dx)/(Area of compliantmember))²×deflection stiffness; the system having a resonant frequencygiven by the formula;system resonant frequency=(system effective stiffness/system effectivemass)^(1/2); wherein the cross sectional area and length of the inertiatrack and the deflection stiffness of the compliant tuning member arechosen such that the system resonant frequency is generally at thechosen frequency; wherein the stiffness of the variable volume chamberat frequencies substantially below the chosen frequency is less than 50N/mm.